The development of computer hard disk drives demands ever increasingly higher track density, lower acoustic noise, and better reliability under shock and vibrational disturbances. The undesirable characteristics of the currently used ball bearing spindles, such as high non-repetitive runout, large acoustic noise, and high resonance frequencies due to bearing defect, impose severe limitation on the drive's capacity and performance.
The use of a non-contact bearing, such as a hydrodynamic bearing ("HDB"), may overcome the aforementioned limitation. The full film lubrication of a fluid bearing displays significantly lower non-repetitive runout and acoustic noise, and its higher damping provides better resistance to external shock and vibration. One example of a disk drive spindle motor including a HDB and centrifugal-capillary seals is found in the present inventors' (with another co-inventor) U.S. Pat. No. 5,423,612 entitled: "Hydrodynamic Bearing and Seal", the disclosure thereof being incorporated herein by reference.
The deployment of the HDB system in a hard disk drive environment requires that the lubricant be securely sealed inside of the bearing structure under all operating and non-operating conditions in order to prevent performance degradation of the bearing and contamination in the drive. At the same time, the bearing system needs to be easily manufacturable in order to satisfy cost requirements. As explained below these requirements often come into conflict with each other and have heretofore resulted in compromised HDB spindle designs.
There have been a number of prior approaches for providing seals for hydrodynamic bearing units. Static seals, such as O-rings, and dynamic clearance seals, such as surface tension or capillary seals, have been employed to seal hydrodynamic bearings.
One prior example is found in Hendler et al. U.S. Pat. No. 3,778,123 entitled: "Liquid Bearing Unit and Seal". In the Hendler et al. approach, a non-wettable liquid, such as mercury, is placed in an annular Vee-groove at an outside boundary of the hydrodynamic bearing system. In addition, a thin film of low vapor pressure vacuum pump oil is provided at an annular gap or space at the end of a journal member in order to retain the mercury seal. A pair of thin barrier films are also provided at the outer edge of the annular space to prevent the oil from spreading as a result of surface effects and/or centrifugal forces generated by relative rotation of the bearing system.
Another prior approach is found in Van Roemburg U.S. Pat. No. 4,596,474, entitled: "Bearing System Comprising Two Facing Hydrodynamic Bearings". In the Van Roemburg approach, two radial fluid bearings were separated by a central reservoir. Each bearing included a herringbone pattern, and the herringbone patterns were such that the outer legs of the Vee-grooves forming the herringbone pattern were longer than the inner legs. However, the system maintained balanced pressure. This arrangement built up a lubricating liquid pressure at the apex of each Vee-groove which was greater than a counter pressure built up by the inner legs and by helical feed grooves which feed lubricant from a central reservoir area. By providing this differential pressure arrangement it is said that the lubricant was not pumped out of the bearing system.
A further prior approach is described in Anderson et al. U.S. Pat. No. 4,726,693, entitled: "Precision Hydrodynamic Bearing". The Anderson et al. approach uses a plurality of seals formed along the bearing unit including spiral grooves as well as an upper surface tension or capillary seal and a lower surface tension or capillary seal. However, the very nature of the Anderson et al. approach suggested that it was not adapted to omnidirectional operation or resistance to shock or vibratory forces.
Another prior approach is described in Titcomb et al. U.S. Pat. No. 4,795,275 and divisional patents U.S. Pat. Nos. 5,067,528 and 5,112,142, entitled: "Hydrodynamic Bearing". In the prior approaches described in these patents, surface tension dynamic seals were provided between axially extending surfaces of a thrust plate and bearing sleeve (or between tapered bearing surfaces). Pressure equalization ports were required and extended between the dynamic seals and interior lubricant reservoirs (or interior dynamic seals) to balance the hydrodynamic pressures in the lubricant in order to prevent the lubricant from being pumped through one of the dynamic seals. A method for introducing lubricating liquid into the hydrodynamic bearing employing a vacuum chamber and ultrasound is also described.
A similar prior approach is described in Pan U.S. Pat. No. 5,246,294 entitled: "Flow-Regulating Hydrodynamic Bearing". In this approach a disk spindle employs oppositely facing conical hydrodynamic bearing surfaces and a series of chambers and passages and a gravitational valve are provided to permit pressure-equalized centrifugally pumped global circulation of lubricating liquid drawn from one or more large reservoir volumes. A leak-preventing capillary trap "of minimum continuous axial length" may be provided at a clearance seal for passive capture of wandering lubricant when the bearing unit is at rest.
A further solution has been proposed by the present inventors with another in U.S. Pat. No. 5,423,612 entitled: "Hydrodynamic Bearing and Seal", the disclosure thereof being incorporated herein by reference. One drawback of the approach described in the '612 patent is it has proven somewhat difficult to provide recirculation ports around the bearings in order to realize a lubricant recirculation capability in circumstances such as imbalanced pumping and/or shock load. Another drawback was that since both top and bottom seals are at the inside diameter of the HDB unit, any splashed droplets which separate from the lubricant surface may be driven out of the bearing by centrifugal force. In addition, because of the small seal volume available at the HDB unit inside diameter, the lubricant may leak out of the bearing on account of thermal expansion and/or filling volume variations. A further drawback was that press fitting the thrust plate onto the shaft may cause excessive deformation resulting in large variations in bearing clearances and unacceptable hydrodynamic operation.
Small (3.5 inch disk diameter and smaller) form factor disk drives are used in unlimited applications and orientations. Consequently, a hydrodynamic bearing system for a disk spindle in such drives, e.g. having a full Z-dimension 1.6 inch height spindle manifesting high inertial loading, must also operate in all possible orientations, and to be able to withstand and sustain certain shock events and vibration levels without leakage. A cover-secured or top-fixed HDB motor is required for disk drives with high inertial load, such as disk drives including six or more rotating disks. For top-fixed spindles, the requirement for two lubricant seals poses a considerable challenge.
Generally, there are two types of top-fixed HDB spindle designs, namely: single thrust-plate design, as illustrated in commonly assigned U.S. Pat. No. 5,423,612; and, double thrust-plate design, as illustrated in FIG. 1.
The single thrust-plate design of the type illustrated in U.S. Pat. No. 5,423,612 has the several drawbacks already noted above.
The double thrust-plate design is shown in FIG. 1, and it overcomes the first and second drawbacks of the conventional single thrust plate designs noted above, while the third drawback remains present. Unfortunately, there are some additional drawbacks. As shown in FIG. 1 a prior double thrust-plate hydrodynamic spindle bearing system 10 for a high performance miniature hard disk drive includes a base 12 and a shaft 14 fixed securely to the base 12 in a suitably sized opening 13 defined in the base 12. A shaft housing 16 fits closely over the shaft 14 and cooperatively defines two hydrodynamic journal bearings 34 and 36. A spindle hub 18 is attached to the shaft housing 16 and a flange 24 of hub 18 supports one or more data storage disks 20 (and spacers 22) in a top-fixed arrangement(see e.g. FIG. 2). An in-hub spindle motor rotates the hub 18 and disks 20 relative to the base 12 and shaft 14 at a predetermined angular velocity.
An upper annular thrust bearing plate 28 fits securely over the shaft 14, while a lower annular thrust bearing ring 30 also fits securely over the shaft 14. Together, the plate 28 and ring 30 cooperate with adjacently facing radial faces of the shaft housing 16 to provide an upper hydrodynamic axial thrust bearing 40 and a lower hydrodynamic axial thrust bearing 42. A central axial reservoir region 38 is provided for lubricating liquid between the two radial hydrodynamic bearings 34 and 36. Two end reservoirs 44 and 46 are formed respectively between the bearings 34 and 40, and the bearings 36 and 42.
Two surface tension annular capillary seals 48 and 50 are provided in annular gaps outwardly beyond the two thrust bearings 40 and 42 relative to the shaft 14. The upper seal 48 is formed by outwardly axially divergent, oppositely facing cylindrical walls of the thrust plate 28 and shaft housing 16, and the lower seal 50 is formed by outwardly axially divergent, oppositely facing cylindrical walls of the thrust ring 30 and the shaft housing 16. In these seals 48 and 50, a curved lubricant-air interface typical of a surface tension interface is located approximately midway of the gap.
Two oil containment bushings 52 and 54 are secured in a sealed arrangement to the shaft housing 16 as shown in FIG. 1. The seal 48 is shown as having a radially inward extension 49, which effectively extends the upper capillary seal. The lower capillary seal 50 is likewise extended radially inwardly by an extension 51. Further details of this prior arrangement are disclosed in commonly assigned, copending U.S. patent application Ser. No. 08/363,566 filed on Dec. 22, 1994, entitled: "A Self-Contained Hydrodynamic Bearing Unit" now U.S. Pat. No. 5,558,445, the disclosure thereof being incorporated herein by reference.
As noted above, there are several additional drawbacks with the double thrust-plate design illustrated in FIG. 1. One additional drawback is controlling the tolerance of the total length of the sleeve which defines the thrust bearing clearance (which is about 10 microns). A second drawback relates to manufacturing difficulty in controlling the tolerances of perpendicularity and surface finish at both ends of the e.g. bronze sleeve. Third, because the sleeve is typically of bronze, the sleeve tends to wear by coming into contact with the grooved steel thrust plate having pumping grooves during starting and stopping intervals. Fourth, it has proven difficult to apply adhesive to seal the thrust plate/shaft press-fit areas of the bearing unit. Adhesive grooves at the side of the bottom thrust plate cause a lack of symmetry in the thrust plate and additional deformation during press fitting of the thrust plate and shaft.
Thus, a hitherto unsolved need has remained for a hydrodynamic bearing system having a high inertial load which is leak free irrespective of orientation, shock and vibration, and which is readily and reliably manufacturable at reasonably low manufacturing cost.